Apparatus for generating and transmitting sonic vibrations



Oct. 20, 1964 A. G. BODINE 3,

APPARATUS FOR GENERATING AND TRANSMITTING SONIC VIBRATIONS Original Filed July 6, 1959 3 Sheets-Sheet 1 N Q N w a: Q

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Oct. 20, 1964 A. G. BODINE APPARATUS FOR GENERATING AND TRANSMITTING SONIC VIBRATIONS Original Filed July 6, 1959 $5 Sheets-Sheet 2 INVENTOR.

AL 85/?7' GI Boa/NE Oct. 20, 1964 A. BODINE APPARATUS FOR GENERATING AND TRANSMITTING some VIBRATIONS 3 Sheets-Sheet 5 Original Filed July 6, 1959 E M m a a United States Patent 3,153,530 APPARATUS FOR GENERATEJG AND TRANS- MlTTlNG SONIC VKBRATIONS Albert G. Bodine, Sherman Oaks, Calif. (7877 Woodley Ave, Van Nuys, Calif.)

Original application July 6, 1959, Ser. No. 825,117, new Patent No. 2,968,314, dated Nov. 15, 1960. Divided and this application Sept. 12, 1960, Ser. No. 55,537

Claims. (Cl. 259-4) This invention relates generally to apparatus for the generation and transmission of relatively high power by means of intense sonic vibrations, particularly for generation or transmission of sonic vibrations in resonant vibratory mechanical devices, either elastically deformable vibratory bodies of the distributed constant class, or elas tically supported bodily vibratory devices of lumped constant characteristics.

This application is a division of my application entitled Method and Apparatus for Generating and Transmitting Sonic Vibrations, Serial No. 825,117, filed July 6, 1959, now Patent No. 2,960,314. Said application Serial No. 825,117 was a continuation-in-part of my prior applications as follows: Serial No. 313,175, filed October 4, 1952, for Method and Apparatus for Generating and Transmitting Sonic Vibrations, allowed April 14, 1958, now abandoned; Serial No. 771,808, filed September 2, 1947, entitled Apparatus for Boundary Layer Control, now abandoned; Serial No. 484,627, filed January 28, 1955, entitled Apparatus for Generating and Transmitting Sonic Vibrations, now abandoned; Serial No. 572,432,

led March 19, 1956, for Sonic Materials Separation Process and Apparatus, now abandoned; and Serial No. 772,270, filed November 6, 1958, for Method and Apparatus for Generating and Transmitting Sonic Vibrations, now Patent No. 2,960,317. 7

A large number of industrial uses for high power sonic vibrations have been discovered. It is known, for example that intense sonic energy may be applied to gases, liquids or solids to produce certain desired chemical or physical effects. Many types of power tools or other equipment are operated by sonic energy of high intensity. One illustrative example involves a longitudinally eX- tended elastic bar, in which a longitudinal resonant standing wave is set up and maintained, so that an end portion of the bar becomes the location of a velocity antinode of such standing wave, and is utilized to vibrate a bit or other tool against the Work. Other modes of vibration, such as lateral or gyratory, are within the scope of the invention. Other illustrative examples will appear in the ensuing description.

The bodies or devices to be sonically vibrated at resonance are often characterized by high acoustic impedance. They vibrate with great force, and with small velocity amplitude. The problem of driving these devices, i.e., the provision of an effective vibration generator suited thereto, is often very diiiicult, particularly in view of the fact that ordinary practically available sources of motivating power operate at low impedance, characterized by driver elements moving with relatively low force but substantial velocity. Ordinary low impedance drivers are incapable of driving high impedance devices because of the mismatch of impedance. The efiiciency of transduction has been characteristically low.

The common sonic generators, such as magneto-striction bars, crystals, etc., are capable of a motion of only a few feet per second by reason of limitations set by elastic strain limits, which motion I have found to be entirely inadequate for high power applications.

Mechanical generators are known which have the requisite motional characteristics, but suffer from complexity and a host of ensuing problems. Any degree of complexity of moving parts results in various vibratory interactions taking place at high frequency between these parts, with consequent high energy loss and frequent destruction of parts in high stress applications. At very high frequencies gears will chatter, bearing separators seize and fracture, and the individual balls or rollers of anti-friction bearings are forced to rotate so fast they ,become unstable in their motion. Plain journal bearings seize and overheat. The power of previously known generators has been relatively low, particularly at the higher frequencies; and the ruggedness required of an industrial machine has been lacking. Many proposed industrial applications of sonic power have been correspondingly handicapped.

It is accordingly the primary object of the present invention to provide novel and improved sonic vibration generating apparatus particularly suited to various industrial applications and characterized by relatively high power output, efiiciency of transduction, simplicity and ruggedness.

The invention is practiced in systems involving the driving of an inertia mass rotor in an orbital path under guiding constraint of a bearing means, whereby a periodic force impulse is exerted on the latter, and the coupling of this hearing means to a vibratory device having a resonant frequency range whereby said periodic force impulse, or a component thereof, is effective to vibrate said device in said range. To this end, the rotor is driven at an orbital frequency which generates a vibration frequency in the range of resonance for the driven vibratory device. I have discovered that the driven vibratory device, when so vibrating in its resonance range, with its vi bration amplitude amplified by resonance, back-reacts with the orbital rotor, strongly constraining the rotor to an orbital periodicity corresponding to its own resonant frequency. I have further discovered that the apparatus tends inherently to operate on the low side of the frequency for peak resonant amplitude, and further, that the whole apparatus, driven vibratory device and orbiting rotor, tends to lock in synchronously slightly below the frequency for peak resonant amplitude. The orbiting rotor is strongly constrained to produce this frequency, and although it could of course be strongly enough driven to reach a threshold condition where it would reach and break over peak amplitude resonance frequency, considerable increase in driving effort is required before this unwanted condition occurs. In this connection, it is to be understood that the driving effort on the rotor is limited to a value below such threshold condition. The rotor is hence guarded from over-speeding and destroying itself or its housing when operated at high frequency.

In addition to these effects, the constraint which keeps the frequency of the orbiting rotor to the low side of the resonance curve (amplitude vs. frequency) of the vibratory driven device is effective to establish a phase angle between the rotor motion and the motion of the vibrating device wherein maximum power is delivered from the rotor to the vibrating device for a given power input to the rotor.

It will be evident that such an orbiting rotor generator has high output impedance, while being operable by motive power at low impedance, impedance being understood to be proportional to the ratio of force to velocity. Considering the output side of the generator, where the race for the orbiting rotor is coupled to the vibratory driven device, it will be seen that force will be high owing to the high magnitude of centrifugal force, while the stroke amplitude, and therefore the velocity amplitude, will obviously be low. The desirable high output impedance for the resonant system is therefore attained.

Patented Oct, 20,

Impedance is generally thought of in connection with alternating phenomena such as alternating forces, in comparison with resulting velocity amplitude. The motive power source used in the present instance is typically a continuous air jet, rather than an alternating entity. Nevertheless, the continuous air jets has the characteristic of relatively low force and relatively high velocity, and is, broadly speaking, a form of power having a low impedance quality. The generator of the invention thus fulfills the requirement of operating off a low impedance form of power, and delivering power at high impedance.

When considering the orbital path it is important that the well-known laws of acceleration be recognized, because it is the phenomena of acceleration which necessarily causes the periodic force generation by the periodic movement of the mass. Acceleration is inherently the most important phenonmena involved in orbital movement. Acceleration basically consists of the phenomena of change of velocity. Change of velocity, of course, involves change of direction, or change of speed, or a combination of both. Most practical examples of acceleration involve both change of direction and change of speed. Most forms of this invention involve this combined form of acceleration, as will be evident from a study of the examples cited.

The authorities consider the various classical examples of orbital paths to be forms of ellipses. In this regard a circle is considered one form of ellipse. With a circular orbit the acceleration is usually a uniform change of direction, with no change of speed. Ellipses can have various ratios of major to minor axis, the limiting wellknown cases being the circle as one limit and practically a straight line of linear reciprocation as the other limit. In any noncircular ellipse the all-important acceleration phenomena is concentrated at the two ends of the pattern. Here we have a combination of change in direction and in speed. In many of the examples described herein the orbital motion of the mass is necessarily some form of noncircular ellipse path, as above described. In many cases the shape of the path in space changes during operation, such as from changes in work load, being influenced by the vibration pattern of the resonant member to which it is coupled.

As will be evident from a study of the figures, some of the resonant members require that the resultant of the orbital excitation be circular; and some require that the resultant of the orbital path be linear. The cyclic (orbital) motion of the moving mass need not in all combinations correspond exactly with the shape of vibratory motion path carried on by the elastic member. The elastic member in some cases shown herein has a definite resonant response consisting of linear vibration. In such cases the excitation orbit can, of course, also be linear in effect by any convenient means, or it may be a more circular ellipse. In the latter such cases the sharpness of tuning of the resonant member will cause it to vibrate predominantly in a straight line, and will cause the mass orbit to become predominantly ellipsoidal sometimes substantially a straight line, even when the mass orbit by itself be very circular. As can be seen it is only necessary that the orbiting mass and the resonant member be interconpled with a common component of cylic force in order to be able to have the back-reaction eifect which is the key to this invention. Obviously either the resonant member or the accelerating excitation mass in the novel combination can have any of the classical forms of closed cyclic (orbiting) path, depending upon the needs of the situation. The guidance and path limiting means can thus be arranged to cause the inertia mass to be constrained to any shape of closed cyclic path, including circles, straight lines, rippled or carnmed circles, camrned ellipses, various other ellipses, etc., so long as the path limiting means receives the impulses from the mass and is coupled to the resonant means in vibration transmitting relationship so that the acceleration impulses are transmitted thereto.

The illustrative embodiments chosen for disclosure herein are for the most part of the type wherein the resonantly driven device is of the distributed constant type. though without implied limitation thereto. It should be explained that a distributed constant system is one wherein the parameters of mass and elasticity governing the resonant vibration frequency are distributed throughout all or a significant part of the vibratory system, as in the ideal example of a tuning fork. By contrast, a lumped constant system is one wherein the parameters of mass and elasticity governing the resonant frequency are largely concentrated or localized in discrete elements such as intercoupled masses and springs, respectively. Of course, these are idealized classifications. Practical systems usually are mixtures of the two. Thus, practical systems wherein the parameters of mass and elasticity are preponderan-tly distributed will also very commonly have local concentrations of mass, with small capability for elastic vibration therein; while practical systems wherein mass and elasticity are preponderantly localized, as in intercoupled spring and mass elements, will invariably have certain distributed constant qualities in view of mass inherently present in spring elements, and elasticity inherently present in mass elements. Thus, the resonantly driven devices of the invention may embody such distributed constant elements as an elastic bar, in which either transverse, gyratory, or longitudinal standing wave action may be set up by the vibration generator. Such bar may be a solid rod, or it may be tubular, as a steel pipe. The term bar is often used in the field of acoustics in connection with discussions of elastic wave propagation, without reference to the cross sectional form of the bar, and the term will be so used herein, both in the specification and claims, thus generically comprehending hollow rods, or pipes, as well as solid rods, I-beams, and other structural shapes.

One aspect of the invention is based on the fact that an elastic bar, properly supported, can be subjected to transverse vibration simultaneously in two directions at right angles, that is, in a manner such that a free end of the bar describes a closed path, which is the resultant of two rectilinear transverse components of vibration. If the bar is of circular section, and the two components are in quadrature and of equal amplitude, the free end describes a circle. If the two components in quadrature are of unequal amplitude, the free end will describe an ellipse. It can readily be seen that the effect of the two quadrature vibratory components is a rotating radially directed force vector which bends the portion of the bar acted upon in a circle, but, of course, without turning the bar bodily on its axis. The advantage of so setting up a rotating or gyratory deflection of the bar at the point of vibration generation, as compared with a simple unidirectionally transverse vibration, is that for a given amplitude of lateral deflection, a bar so vibrating receives, transmits and delivers twice as much sonic power as in the case of unidirectional transverse vibrations.

Transverse gyratory vibration of an elastic bar of the character indicated results in propagation of an elastic gyratory transverse wave in a direction longitudinally of the elastic bar, each transverse section of the bar remaining in its own plane, and successive transverse sections undergoing gyratory transverse elastic deflection. It can be seen that such elastic wave propagation longitudinally of the bar permits vibratory energy to be taken oil from the end of the bar remote from the end at which the vibrations are imparted to the bar.

The orbiting rotor generator of the invention has especially useful properties when coupled to drive the gyratory bar of the system described in the immediately preceding paragraphs, as will appear.

The invention will be further described in connection with the following detailed description of a number of illustrative embodiments thereof, reference being had to the accompanying drawings, in which:

FIG. 1 is a longitudinal sectional view of one embodiment of the invention;

FIG. 2 is a transverse section taken on line 22 of FIG. 1;

FIG. 3 is a section taken on line 33 of FIG. 2;

FIG. 4 is a diagrammatic view illustrative of a standing wave action characteristic of the apparatus of FIG. 1;

FIG. 5 is a side elevation, partly in medial section, of a modified form of the invention;

FIG. 6 is a section on line 66 of FIG. 5;

FIG. 7 is a section on line 77 of FIG. 6;

FIG. 8 is a detail side elevation of the rotor of FIGS. 5 and 6;

FIG. 9 is a longitudinal medial section through a vibration generator in accordance with the invention, the generator being illustrated as formed on the extremity of a fragmentarily illustrated elastic bar of the type more fully shown in FIG. 1;

FIG. 10 is a view similar to FIG. 9 but showing a modification;

FIG. 11 is a view similar to FIG. 9 but showing a modification; and

FIG. 12 is a view similar to FIG. 9 but showing a modification.

FIGS. 1 to 4 illustrate a system wherein the invention claimed herein may be applied. The case illustrated involves the problem of supplying intense sonic energy to liquids or gases to produce desired physical or chemical effects. Many industrial uses for such a process are known and described in the literature and need not be further discussed herein.

The elastic bar is here in the form of a tube 20, typically composed of steel, which is carried by spaced rubber blocks or sleeves 21 supported by mountings 22, and these blocks 21 are such as will permit a substantial degree of elastic vibration in all directions in planes transverse of the tube. The tube does not rotate bodily, but portions thereof spaced from the nodal point or points of a standing wave set up in the tube gyrate in a circular path by elastic bending of portions of the tube from its neutral position (see FIG. 4). Such circular motion or gyration is a form of harmonic vibration, being the resultant of two components of linear transverse harmonic vibration occurring at right angles to one another with 90 phase difference. The rubber blocks 21 will be seen to comprise compliant mountings permitting such gyratory action.

The vibration generator 24 here shown is of a gyratory ring type, but could equally have been of a roller type shown in FIGS. 5-9 of my Patent No. 2,960,314. It comprises a cylindric housing 25 formed with a cylindric chamber 26, preferably though not necessarily, co-axial with the tube 20. This housing 25 is formed with one integral side closure wall 25a, and its opposite side is fitted with a removal closure wall 27. A flanged fitting 28 is secured to the wall 25a, and has a threaded projection 29 screwed into the corresponding end of tube 20. A center pin or axle 30 of circular cross section, preferably formed with a central crowned or barrelshaped portion 31, has reduced end portions 32 set tightly into the walls 25a and 27. The periphery of this axle 3% provides a circular rolling bearing surface, which is surrounded by an inertia rotor or roller in the form of a ring 33, having a circular, smooth-surfaced central opening 34 of substantially larger diameter than that of pin 30. In some applications the outer periphery of the ring has a small clearance with the periphery of the cavity 26 when hanging on the axle 30, or spinning thereabout.

The inertia ring 33 is caused to roll on its axle 31 by a fluid jet, either air under pressure, steam, or a liquid, introduced through an injection nozzle 35 formed in the housing 25 tangential to the periphery of the circular cavity 26, such fluid being introduced to the nozzle 35 via a hose 36 coupled thereto. The spent driving fluid may be discharged from the chamber 26 in any desired manner; as here shown, it is vented to atmosphere via orifices 37 formed in closure plate 27 as close to the center of the chamber 26 as possible.

The tangentially introduced fluid causes the inertia ring 33 to roll on the axle 30, and the centrifugal force exerted by the rolling ring on the axle 3t), and thence transmitted to the housing 25, elastically bends the proximate end portion of the tube 20 and moves it around in a circular path. As earlier pointed out, this motion of the end portion of the tube is a form of harmonic vibration, being the resultant of two perpendicular transverse linear harmonic vibrations in quadrature.

FIG. 4 shows with some exaggeration, the tube 20 undergoing gyratory elastic motion characteristic of a standing wave of the fundamental resonant frequency of the tube for longitudinally propagated transverse elastic Waves. It will be understood from known principles that the standing wave diagrammatically indicated in FIG. 4 results from the transmission down the length of the tube, from the generator 24, of transversely oriented elastic deformation waves, which are reflected from the far end of the tube, and through interference with a succeeding forwardly propagated wave, the standing wave is established somewhat as indicated. It will be seen that nodal points occur at sections of the tube approximately onefourth the length of the tube from each of its ends, whiie.

the two ends of the tube are at antinodes of the standing wave.

The speed of rotation of the inertia ring 33 about the axle 30 is in the first instance determined by the fluid jet which drives it. I have discovered, however, that presupposing a drive of the inertia ring at a number of revolutions per second about the axle 34 which approaches or approximates the resonant frequency of the tube 20 for the described transverse mode of Vibration so that the tube 20 and the gyrator housing 24 connected to one end thereof will describe circles of augmented amplitude at the antinodal points, the inertia ring 33 then unexpectedly tends strongly to lock in at that frequency, i.e., to spin at a number of cycles per second equal to the resonant frequency for the tube 263 and housing 24. I have further found that the spin speed of the ring tends to lag slightly behind the precise resonant frequency for peak resonant amplitude, or in other words, stays on the low side of the resonance curve. Under these conditions, any tendency for overspeeding of the inertia ring with increased pressure on its driving jet is strongly resisted. The resonantly gyrating tube thus exerts a back reaction on the inertia ring, holding the ring at resonant periodicity, but on the low side of the frequency for peak amplitude, thereby preventing it from over-speeding. In other words, under the constraint imposed by the described back reaction from the resonantly vibrating elastic member, the spin frequency of the ring is held to the resonant frequency of the elastic member, slipping to a degree within the driving fluid stream, the fluid jet thus acting as a slip-type drive.

The apparatus of FIGS. 12 is here shown as equipped with means for introducing a fluid to one end of the tube and discharging it from the other. Thus, an inlet tube 48 coupled to a passageway 41 in member 29 introduces the fluid to be treated to one end of tube 20 and an outlet tube 42 mounted in a plug 43 screwed into the opposite end of the tube 20 communicates via a passageway 44 with tube 2t to withdraw treated fluid. It will be understood that fluid within the tube 2i] is subjected to sonic agitation. Various known industrial processes capable of making use of such sonic frequency agitation of liquids or gases form no part of the present invention and need not be described herein.

FIGS. 5 to 8 show a sonic vibration generator utilizing an orbital rotor of the fluid-driven, whirling ring type, the generator being illustratively shown as coupled to an 7 end portion of an elastic bar whose opposite end is cou pled to a work load, in an arrangement such that a longitudinal standing wave is set up in the bar.

In FIGS. and 6, the elastic bar is designated generally at 16d. This bar is understood to be composed of some good elastic material such as a good grade of steel or alloy steel. The upper end of this bar 16% is flangeconnected, as by bolts, or silver soldering, to the diagrammatically illustrated work load Ml, which may be the lower, vibratory wall of a liquid tank whose contents are to be sonically treated, or may be any other device to which vibratory action is to be applied. The lower end of bar 16% is formed with a cylindrical hub 152, into whose bore 163 is press fitted the sonic vibration generator generally designated at 164. Two circular or disklike side plates 165 are press fitted into bore 163 at a spacing to provide a cylindrical chamber 166 for the presently described rotor 167. The side plates ZbSS are hollowed out to provide manifold cavities i655, and tightly mount an axial bearing pin 169 which extends across rotor chamber 166.

Rotor 167, which can be fabricated of ball bearing steel, comprises a relatively thin hub 17% receiving pin 169 with a clearance of the typical proportions shown. Integral with hub 17%) is a medial, radially extending web 171, on opposite sides of which are impeller vanes 172, formed integrally with hub 176i and web 171. These vanes 172 are curved, as shown best in EEG. 24 and are of substantial area, preferably projecting laterally beyond the ends of hub 17%. The vanes are formed with radial, outside edges 173 parallel and very closely spaced to the inside surfaces of plates 1165, the clearance being preferably of the order of .001 inch on each side, so as to prevent the rotor wobbling or otherwise becoming unstable in its high speed gyration. I have also found it desirable to provide a very high polish on the inside surfaces of the side plates so as to reduce friction between the rotor and these surfaces to a minimum.

A plurality of nozzle bores 176 are drilled through the inside walls of plates 165 between manifold chambers 168 and rotor chamber 166. These are preferably placed in a circular pattern around pin lid?) in a pattern such as indicated in FIG. 23, and are oriented in a tangential direction with reference to rotor chamber 166. They are also preferably so positioned that the fluid jets delivered therefrom impinge on the vanes 172, as near as possible to their innermost junction with the rotor hub, so as to cause the air streams to flow radially outward along the vanes as they deliver energy thereto. It will be evident that this desirable condition can be more fully realized than as shown in FIG. 6, if the rotor hub has less clearance with the pin 169; and excellent performance has been attained in practice with very considerably less clearance proportions than as shown in the illlustrative embodiment. The spent air flows off the peripheries of the rotor vanes, and is exhausted via an ample discharge port 177 formed in the pen'phery of hub 162, as shown best in FIG. 6.

The air manifold chambers 168 are fed via a bore 172% in bar 160 leading from an air inlet to a pair of branch passages 179 communicating with ports 180 in side plates 165 opening into cavities 168. A suitable air conduit connection is made to bore 178 through the side of bar 160 as indicated at 181.

The device thus described operates in accordance with principles and in a general manner discussed hereabove in connection with FIGS. 1-4.

It will be observed that in the device of FIGS. 5 to 8, a longitudinal resonant standing wave is set up in the bar 160 with use of but a single orbiting rotor generator, whereas in earlier described embodiments, a plurality of generators were employed in order to cancel out components of vibration at right angles to the direction of the bar. It is found in practice, however, that substantial vibration amplitude in the bar 166* can be made to 8 occur only at the longitudinal resonant mode of vibration of the bar by selection of rotor size to give a powerful impulse only up in this range, and in the absence of capability for vibration in strong lateral modes at the frequency of the longitudinal mode, the component of force delivered laterally by the rotor causes very little lateral vibration.

The invention further provides a means for reducing or virtually eliminating any tendency for permature resonant lock in at unwanted modes of vibration below the desired operating frequency, such as lateral modes, or resonant bouncing modes against the load wherein the bar vibrates bodily. It will be evident to those skilled in the art that the bar 16% of FIG. 5 would have a lateral mode of vibration with a resonant frequency lower than its first longiudinal mode. It would be undesirable to permit such a lateral mode to take over control of the rotor, holding back on its ability to climb in frequency up to the desired longitudinal mode. The rotor could be driven hard enough to pass the frequency of the unwanted lateral mode; but it is preferable to suppress vibration at any such mode.

To prevent or suppress a lateral mode of such strength, I incorporate vibration damping material at a strategic location relative to the bar. For example, and as shown in FIG. 5, l provide the bar 16d in its central region with a longitudinal slot 185, and place therein a body 186 of viscous, pliable damping material, such as tar, pitch, a thermoplastic, or partially vulcanized rubber. The material introduces shear viscosity damping under the conditions of bending with any tendency for lateral vibration bending modes, so as to prevent any substantial tendency for resonant lock-in at such a mode.

To prevent unwanted low frequency longitudinal modes of vibration, such as a resonant bouncing mode wherein the bar vibrates as a whole againt the spring action of the load, I preferably surround the bar, in the region of a node of the desired Wave pattern (which is the longitudinal center point of the bar in the case of FIG. 5), with an inertia mass ring 137, and I interpose between this ring 187 and the bar a ring-like body 188 of damping material of the same viscous, pliable nature as the already described damping body 186. This damping material, in combination with the inertia ring 187, clamps out all longitudinal resonant modes excepting the mode which locates a node of the wave pattern at the damping body.

FIGS. 912 show forms of the invention characterized by vibration generators wherein either the rotor or its bearing has, instead of a circular contour, a cam-shaped contour designed to introduce frequency multiplication. These operate in accordance with the theory and principles of the invention discussed in the introductory part of this specification, and in connection with the devices of FIGS. 1-4 and 5-8, particularly as regards the resonant frequency range of the elastic resonant vibratory member and the driving force exerted on the rotor.

In FIG. 9, a circular bearing 524 is formed on the end of a fragmentarily illustnated elastically vibratory bar 521 having a resonant frequency range, which bar may if desired, be similar to the elastically vibratory resonant bar of FIG. 5. The rotor 522 in this case is formed as a roller with a plurality of arcuate surfaces 523 whose radius of curvature R is relatively long, but less than that of the circular bearing surface of member 520, connected by arcuate surfaces or lobes 524 of substantially smaller radius of curvature R. The rotor may be driven in an orbital path, guided by bearing 520, by air under pressure injected into circular rotor cavity 525 through tangential air nozzle 526. The rotor cavity will be understood to have side wall surfaces which fairly closely confine the rotor, for example, as in FIG. 6, or as in FIG. 5 of my aforementioned Patent No. 2,960,314. The described contour causes the rotor to have a periodic component of motion radial of the axis of the orbital path, i.e., radial of the bearing as well as to roll around said bearing, and thereby to generate an additional frequency which is equal to the number of lobes times the orbital spin frequency of the rotor. It is important that the cam contours be designed so that the rotor is not radially accelerated thereby sufliciently to overcome the centrifugal force which holds the rotors in contact with the bearing, since otherwise the rotor would separate period ically from the bearing surface, and impact thereagainst on its return.

In FIG. 10 which is a reversal of FIG. 9, a bearing 527 on a vibratory bar 528, like bar 521 of FIG. 9, has arcuaite surfaces 529 of relatively long radius of curvature, connected by arcuate surfaces 530 of relatively smaller radius of curvature, as indicated, the rotor 531 being round (ball-shaped or cylindrical), and of radius of curvature somewhat larger than that of surfaces 536). The drive is again by means of a tangential air nozzle 552. The effect is again the generation of a secondary frequency, which is a multiple of the orbital spin frequency of the rotor.

In FIG. 11 a vibratory bar 533 has at one end a bearing head 534 formed with a cylindrical cavity 535 and provided with an axial pin 536 extending thereacross, which is surrounded by a gyratory ring 537, :of a nature disclosed in earlier embodiments of the invention, for example, FIGS. 13. The pin or axle 5336 is contoured like the rotor 522 of FIG. 9, as indicated by the radii in dicated in the figure, with corresponding generation of a secondary frequency when the ring is spun on the axle by air injected at 538.

Finally, in FIG. 12, which is a reversal of FIG. 11, a vibratory bar 54% has a bearing head formed with a cavity 541, as in FIG. 11, and, extending across, an axial pin 542 which, in this case, is of circular contour, in the arrangement of FIGS. 1-3. The gyratory ring 543 in this case has interior surfaces 544 of relatively long radius of curvature, connected by surfaces 545 of relatively short radius of curvature, the radius of curvature of the latter surfaces necessarily being greater than that of the pin 542. The ring is again by air injected as at 54-6. Again a secondary frequency is imposedon the primary or orbital spin frequency.

The consideration mentioned in connection with FIG. 9 as regards avoidance of cam contours which would cause separation and impacting between the rotor and hearing will be understood to apply equally to the embodiments of FIGS. 10-12.

In all the embodiments of FIGS. 9-12, the primary orbital spin frequency can be damped out, and the higher secondary or multiple frequency permitted to predominate and control simply by utilization of vibration damp ing means of the type disclosed at m7 and 188 in FIG. 5. Thus the vibratory bar can be designed for standing wave vibration (e.g. at half Wave length) at the secondary frequency, and the lower frequency primary or orbital spin frequency, which may tend to cause vibration of the bar as a whole, can then be damped by location of a means such as 187, 188 of FIG. at a node of the desired secondary frequency. The frequency generated, particularly when the orbital spin frequency is damped out in the driven device, will be seen to be that generated by the orbiting rotor owing to its radial component of motion, rather than the orbital spin frequency.

It will be appreciated that the various forms of vibratory devices of the invention will have one or more resonant frequencies or resonant frequency ranges, such as fundamental, or other frequency modes: and itis to be understood that, within the broad scope of the invention, the resonant frequency or resonant frequency range chosen in any particular application may be either the fundamental, or any other desirable frequency mode. The important consideration is that a resonant frequency be used.

It will be clear that any of the generators of FIGS. 9l2 may be substituted for the generator 24 of FIGS. 1-4

IQ and be thus utilized to set up a gyratory standing wave in the tube 26) of FIG. 1, but with frequency the multiplication feature which characterizes the generators of FIGS. 9-12.

The invention has now been described in various forms and applications, which are illustrative and from which its broad scope may be understood. Moreover, it can be seen that many of the features illustrated for convenience in certain embodiments apply equally well to other embodiments. It is of course to be further understood that these are not limitative or a restriction on the scope of the broad invention, excepting as may be fairly construed from the following claims.

I claim:

1. In sonic vibration processing means, the combination of: an apparatus having a vibratory work-delivering part in combination with an elastic member having a resonant frequency range and a part free to vibrate elastically in said frequency range, said resonant frequency range being characterized by having a frequency at which the vibration amplitude has a peak maximum value, and sonic vibration generating means for generating vibration in said elastic member in said resonant frequency range thereof comprises a bearing means coupled to said free part of said elastic member to vibrate therewith, an inertia rotor guided by said bearing means for rotating in an orbital path, driving means for driving said inertia rotor around said path to vibrate said elastic member at a frequency in said resonant range, said driving means having a driving force on said rotor which is less than the threshold value which will cause said resonant peak amplitude, so that said inertia rotor assumes and holds a synchronous relation to the elastic vibration of said free part below the frequency for said peak amplitude, and wherein said bearing means is formed with a plurality of cam contours on its surface so as to impart a plurality of vibrations for each cycle of said rotor.

2. In sonic vibration processing means, the combination of: an apparatus having a vibratory Work-delivering part in combination with an elastic member having a resonant frequency range and a part free to vibrate elastically in said frequency range, said resonant frequency range being characterized by having a frequency at which the vibration amplitude has a peak maximum value, and sonic vibration generating means for generating vibration in said elastic member in said resonant frequency range thereof comprised of a bearing means coupled to said free part of said elastic member to vibrate therewith, an inertia rotor guided by said bearing means for rotating in an orbital path, driving means for driving said inertia rotor around said path to vibrate said elastic member at a frequency in said resonant range, said driving means having a driving force on said rotor which is less than the threshold value which will cause said resonant peak amplitude, so that said inertia rotor assumes and holds a synchronous relation to the elastic vibration of said free part below the frequency for said peak amplitude, and wherein at least one of said bearing means and said rotor is formed with a plurality of cam contours on its surface so as to impart a plurality of vibrations for each cycle of said rotor.

3. In sonic vibration processing means, the combination of: an apparatus having a vibratory work-delivering part in combination with an elastic member having a resonant frequency range and a part free to vibrate elastically in said frequency range, said resonant frequency range being characterized by having a frequency at which the vibration amplitude has a peak maximum value, and sonic vibration generating means for generating vibration in said elastic member in said resonant frequency range thereof comprised of a bearing means coupled to said free part of said elastic member to vibrate therewith, an inertia rotor guided by said bearing means for rotating in an orbital path, driving means for driving said inertia rotor around said path to vibrate said elastic member at a frequency in said resonant range, said driving means having a driving force on said rotor which is less than the threshold value which will cause said resonant peak amplitude, so that said inertia rotor assumes and holds a synchronous relation to the elastic vibration of said free part below the frequency for said peak amplitude, and wherein said rotor is formed with a plurality of cam contours on its surface so as to impart a plurality of vibrations for each cycle of said rotor.

4. In a sonic vibrator, a bearing means, an inertia rotor guided by said bearing means for gyration thereabout, and driving means for driving said inertia rotor about said bearing means to vibrate said bearing means, at least one of said bearing means and said rotor being formed with a plurality of cam contours on its surface so as to impart'a plurality of vibrations for each cycle of said rotor.

5. In a sonic vibrator, a bearing means, an inertia rotor guided by said bearing means for gyration thereabout, and driving means for driving said inertia rotor about said bearing means to vibrate said bearing means, said bearing means being formed with a plurality of cam contours on its surface so as to impart a plurality of vibrations for each cycle of said rotor.

6. In a sonic vibrator, a bearing means, an inertia rotor guided by said bearing means for gyration thereabout, and driving means for driving said inertia rotor about said bearing means to vibrate said bearing means, said rotor being formed with a plurality of cam contours on its surface so as to impart a plurality of vibrations for each cycle of said rotor.

7. The subject matter of claim 5, wherein said bearing means comprises the inside surface of a bearing race and said rotor is a roller which rolls around said inside surface.

8. The subject matter of claim 5, wherein said bearing means comprises a pin with said cam contours formed thereon, and said rotor comprises a ring surrounding and gyratory on said cam-contoured pin.

9. The subject matter of claim 6, wherein said bearing means comprises the inside surface of a bearing race and said rotor is a roller which rolls around said inside surface.

10. The subject matter of claim 6, wherein said bearing means comprises a pin and said rotor comprises a cam-contoured ring surrounding and gyratory on said pin.

References Cited in the file of this patent UNITED STATES PATENTS 2,194,410 Svenson Mar. 19, 1940 2,204,472 Caquot June 11, 1940 2,496,291 High Feb. 7, 1950 2,675,777 Lachaise Apr. 20, 1954 2,960,314 Bodine Nov. 15, 1960 

5. IN A SONIC VIBRATOR, A BEARING MEANS, AN INERTIA ROTOR GUIDED BY SAID BEARING MEANS FOR GYRATION THEREABOUT, AND DRIVING MEANS FOR DRIVING SAID INERTIA ROTOR ABOUT SAID BEARING MEANS TO VIBRATE SAID BEARING MEANS, SAID BEARING MEANS BEING FORMED WITH A PLURALITY OF CAM CONTOURS ON ITS SURFACE SO AS TO IMPART A PLURALITY OF VIBRATIONS FOR EACH CYCLE OF SAID ROTOR. 